Control system for controlling the suspension of a land vehicle

ABSTRACT

The invention relates to an active suspension system for controlling the suspension of a wheeled land vehicle, which comprises an actuator (31) of variable length in parallel with a spring (34), connected between the body or chassis of a vehicle and a wheel and hub assembly (40) of such. The extension/contraction of the actuator is controlled by a control unit (60) in response to parameters measured by various sensors (61, 64, 65, 66 and 62). The control unit (60) also makes allowance for loads on the vehicle not measured by the sensors, so that accurate control of the suspension system is achieved.

This invention relates to a control system for controlling thesuspension system of a land vehicle, in particular a motor vehicle. By"land vehicle" is meant all classes of vehicle capable of powered motionover the land, including motor cars, motor cycles, tractors and trackedvehicles.

In particular, the invention relates to a control system for controllinga motor vehicle having an active suspension system.

An active suspension system is a motor vehicle suspension system inwhich the conventional suspension components, such as springs anddampers, are assisted or replaced by actuators operable in response tocommand signals from a microprocessor in order to correct, change orcontrol the attitude of the vehicle. The aim of such a system is tominimise variations in the forces experienced by the vehicle body,thereby improving vehicle safety and enhancing driver and passengercomfort.

The command signals are produced as a result of the measured values of anumber of variables defining the motor vehicle's attitude. In a trulyactive suspension system, there is the capability not merely ofoperating the actuators in response to previously measured changes inthe values of the variables, but also of, for example, biassing theattitude of the vehicle in some way to offset the effects of asteady-state or dynamic loading; and even of operating the actuators ina manner predictive of expected road conditions.

Active suspension systems are now well known. For example, EuropeanPatent Application number EP-A-0114757 discloses an active suspensionsystem for a four-wheeled motor vehicle in which force measurements aretaken at the points of support of the vehicle body on each wheel/hubassembly and processed to produce a demanded output of the actuatorsecured to operate between each wheel/hub assembly and the body of thevehicle.

A control system for controlling the attitude of a motor vehicle havingan active suspension system is known to include means for converting theforces measured at the points of support of the vehicle body or thewheel/hub assemblies to a set of modal forces which act on the vehiclebody as a whole relative to the wheel/hub assemblies. The requiredactuator outputs may then be calculated to overcome the combined modalforces in order to maintain the desired attitude of the vehicle. Themodal forces are the heave, pitch, roll and warp forces.

A significant advantage of an active suspension system is that thesuspension characteristics of a vehicle may be continuously altered toaccommodate varying road conditions and operating conditions of thevehicle. This facility permits the constructions of a vehicle which hasimproved safety characteristics, since it is possible to maintain agreater degree of contact of the vehicle tyres with a road surface, andthe vehicle is likely to be more predictable to the driver, than in thecase of a vehicle not having an active suspension system.

However, a disadvantage of an active suspension system which resolvesthe forces measured at the points of support of the vehicle body on eachwheel/hub assembly into modal forces is that the resulting modal forcevalues do not accurately reflect the net forces experienced by thevehicle body as a result of road inputs. (By "road inputs" is meant theforce effect of the vehicle passing over irregularities in a roadsurface, such irregularities being, for example, bumps and depressions.)

The active suspension systems currently available lack precision sincethey do not take into account all of the loads transmitted from theunsprung matters of the vehicle, the wheel and hub assemblies, to thesprung matter of the vehicle, the chassis or body of the vehicle. Thesuspension linkage arrangement commonly used by vehicles transmits loadfrom the unsprung masses to the sprung mass by a plurality of loadpaths. To date, active suspension system have only considered thoseloads which can be measured, that is to say the loads on the actuatorand spring assembly.

According to the invention, there is therefore provided a control systemfor controlling an active suspension system of a land vehicle,comprising:

means for measuring loads between the sprung mass of said vehicle andone or more unsprung masses connected thereto;

means for producing a plurality of signals which include the values ofmeasured said loads;

means for modifying said signals to compensate for unmeasured loadsbetween the sprung mass and the unsprung masses; and

means for applying forces to control the attitude of said vehicle independence on said modified signals.

According to a second preferred aspect of the invention, there isprovided a control system for controlling an active suspension system ofa land vehicle, comprising:

means for measuring the vertical accelerations of one or more unsprungmasses connected to said vehicle;

means for producing signals proportional to said measured accelerations;

means for modifying said signals selectively to increase or decrease theextent to which each such signal is dependent on the magnitude of itscorresponding unsprung mass; and

means for including said modified signals in the output signals of saidactive suspension system, means being provided to apply forces to saidvehicle to control its attitude in dependence on said output signals.

There now follows a description of a specific embodiment of theinvention, by way of example, with reference being made to theaccompanying drawings in which:

FIG. 1 is a schematic representation of the effects of heave forces onthe body of a motor vehicle not having an active suspension system;

FIG. 2 is a schematic representation of the effects of pitch forces onthe body of a motor vehicle not having an active suspension system;

FIG. 3 is a schematic representation of the effects of roll forces onthe body of a motor vehicle not having an active suspension system;

FIG. 4 is a schematic representation of the effects of warp forces onthe body of a motor vehicle not having an active suspension system;

FIG. 5 is a schematic representation of a portion of a control systemaccording to the invention.

Referring to the drawings, FIGS. 1 to 4 show schematically a sprung massin the form of a motor vehicle body 20 having four associated unsprungmasses, i.e. four wheels 11, 12, 13, 14 and a respective interconnectingsuspension system (not shown) for each wheel. The vehicle body includesthe engine, transmission and all the ancillary components of the motorvehicle. The motor vehicles shown schematically in FIGS. 1 to 4represent the behaviour of known vehicles not having a control systemaccording to the invention.

FIGS. 1 to 4 are schematic representations of the typical displacementsof a motor vehicle body 20 occasioned by heave, pitch, roll and warpforces respectively. In FIGS. 1 to 4, a numbering convention is adoptedsuch that the front, left hand wheel of the vehicle is labelled 11, thefront, right hand wheel is labelled 12 and so on to the rear, right handwheel which is labelled 14; and the heave, pitch, roll and warp forcesare respectively indicated by arrows H, P, R and W in theircorresponding drawing figures. The modal forces shown in FIGS. 1 to 4are drawn acting positively according to the sign convention adopted.The front of the vehicle is indicated generally by the reference numeral21.

In FIG. 1 the modal force of heave is an equal downward force acting onall four suspension points of the vehicle body. The vehicle body 20therefore tends to move uniformly downwardly without tilting in anydirection under the influence of positive heave.

Positive pitch modal forces are illustrated in FIG. 2 and it is clearthat a positive pitch modal force applied to the vehicle body tends toresult in downward displacement of the front end 21 of the body with notilting from side to side, and with the rear 22 of the vehicle beingdisplaced upwardly from its original position.

The modal force of roll is shown in FIG. 3 as tending to produce atilting displacement of the vehicle body about its longitudinal axis.Positive roll forces therefore tend to produce downward displacements ofthe left hand side of the vehicle body and upward displacements of theright hand side.

FIG. 4 shows the effect of a positive warp force on the vehicle body. Awarp force tends to displace one pair of diagonally opposite corners ofthe vehicle body downwardly and the other pair upwardly in the case of agenerally rectangular body. According to the sign convention usedherein, the front left and rear right corners of a vehicle activesuspension system are downwardly displaced for positive values of warpforce.

It is helpful when considering the forces experienced by a vehicle bodywhich may be resolved into modal forces to divide them into threecategories.

The static loads of the vehicle represent the reaction forces requiredto support the mass of the vehicle and its cargo/passenger load when thevehicle is stationary.

The steady state loads on a vehicle in motion are those occasioned bythe values of the variables of vehicle motion, such as steering angle,vehicle speed, vehicle acceleration/deceleration and the like.

Dynamic loads imposed on the vehicle result from movement of the wheeland hub assemblies, the unsprung masses, when encountering bumps, dipsand the like in a road. It is necessary when designing an activesuspension system to compensate for the steady state loads, so that thesuspension system does not deflect under the steady state loads, butinstead responds solely to the road inputs. There is an exception tothis, in that it may be desired that the suspension system reacts toforces applied to the vehicle by reason of its cornering almost so as tomodify the attitude of the vehicle throughout cornering to improve roadholding.

The dynamic loads on a vehicle are those occasioned by "road inputs",which the driver of the vehicle cannot predict. Such dynamic loadsoccur, for example, when a gust of cross-wind influences the vehicle,and when a wheel of the vehicle encounters a bump in the road surface.

Referring to FIG. 5, there is shown a schematic representation of acontrol system according to the invention.

The arrangement of FIG. 5 represents one quarter of the control systemof a four-wheeled vehicle, one unsprung mass in the form of a wheel/hubassembly being shown, the system for the remaining three unsprung massesbeing similar.

In FIG. 5, the sprung mass of the vehicle in the form of the vehiclebody 20 is shown supported on a number of suspension componentsindicated generally by the reference sign 30, which are in turnsupported on a wheel and tyre modelled as an unsprung mass 40 in theform of the wheel/hub assembly supported on a spring 41 and damper 42representing the tyre characteristics.

The suspension components 30 comprise means for applying forces tocontrol the attitude of the vehicle, in the form of an hydraulicactuator 31 shown vertically aligned and secured at its upper end to aload cell 32 which is separated from the vehicle body 20 by an isolator33, which may be, for example, a rubber block. The actuator 31 need notnecessarily be vertically aligned, depending on the space available forsuspension components and the suspension layout adopted. The load cell32 is capable of measuring at least a portion of the loads actingbetween the wheel/hub assembly and the vehicle body and producing asignal proportional to the loads.

A spring 34 is shown connected in parallel with the hydraulic actuator31.

The spring 34 does not control the attitude of the vehicle in the waythat it would in a vehicle having a conventional suspension system. Roadspring 34 serves merely to reduce the power consumption of the controlsystem of the invention by bearing a significant proportion of thestatic load of the vehicle body 20.

Thus, the operation of actuator 31 may take place over a wide range ofdisplacements actually to effect control of the vehicle withoutrequiring an excessive power consumption as would normally be requiredif the actuator were to support the static load of the vehicle body 20in addition to controlling the steady state and dynamic loadingsresulting from dynamic and steady state forces acting on the vehicle.

Since the power consumption of actuator 31 is reduced, by the use ofspring 34, its piston area may be designed to be relatively small,thereby producing a compact device. Further, spring 34 serves as afail-safe device in that it supports the static load of the vehicle body20 in the event of total failure of the control system of the invention.

The input and output ports of the hydraulic actuator 31 are connectedvia hydraulic pipework 31a and 31b to a hydraulic control circuit 50including a suitable supply pump 51. The hydraulic circuit 50 operates,via electrical connection 50', under the command of a microprocessor 60which produces a demanded output of the actuator 31 in response to anumber of measured inputs.

The inputs to the microprocessor 60 are as follows:

Line 61' carries the output of accelerometer 61 measuring the verticalacceleration of the unsprung mass 40;

Line 62' carries the output of linear variable inductive transformer(LVIT) 62 measuring the displacement of actuator 31;

Line 63' carries the output of load cell 32 measuring the forcetransmitted to sprung mass 20 via the suspension components 30;

Line 64' carries the output of accelerometer 64 located near the sprungmass centre of gravity and measuring the sprung mass longitudinalacceleration;

Line 65' carries the output of accelerometer 65 located near the sprungmass centre of gravity and measuring the sprung mass lateralacceleration;

Line 66' carries the output of gyrometer 66 located near the sprung masscentre of gravity and measuring the sprung mass yaw rate (ie rotationalvelocity);

Line 67' carries a vehicle speed signal from measuring means (notshown);

Line 68' carries a steering rack displacement signal from measuringmeans (not shown);

Line 69' carries a hydraulic system pressure signal from measuring means(not shown); and

Line 70' carries swash plate angle signal from measuring means (notshown) located in the pump 51.

The load cell 32 measures the net load acting between the upper end ofactuator 31 and the vehicle body 20. This load is, consequently,representative of the road input to the vehicle in that a force due to awheel of the vehicle encountering a bump or dip in the road is at leastpartly transmitted to the vehicle body via load cell 32. However, theload measured by load cell 32 generally includes spurious forcemeasurements which it is not required to process, and further does notinclude any allowance for elements of the force due to the wheelencountering a bump or dip which are transmitted to the body viaparallel load paths not including the load cell 32 itself.

FIG. 5 shows a load cell 32 which measures forces transmitted to thebody of the vehicle by both the actuator 31 and the spring 34. However,the applicant envisages a system wherein the load cell 32 measures onlyload transmitted to the body by the actuator 32 and not the loadtransmitted by the spring 34. The load transmitted to the body by thespring can be calculated from the displacement of the actuator 31,measured by the LVIT 62.

The microprocessor 60 of FIG. 5 is capable of resolving the forcesmeasured at each of a number of load cells 32 into a plurality of modalforces corresponding to the modes of vehicle displacement describedabove acting on the vehicle body. Clearly, in the case of a four-wheeledvehicle, the number of measured forces at the load cells 32 associatedrespectively with each wheel/hub assembly is four.

It has been found during the development of the invention that if themodal forces calculated from the measured force values above are used tocontrol the attitude of the vehicle, accurate control is not wihtinacceptable limits. Therefore, the dynamic force on the vehicle bodycause by the acceleration of wheel/hub assembly must also be taken intoaccount, and this is done by multiplying the measured value of thewheel/hub assembly 40 by a mass term.

The heave modal force, for example, is therefore isolated in thefollowing expression: ##EQU1## In which: Hf=Generalised heave force

IVrfH=Front inverted heave load velocity ratio

IVrrH=Rear inverted heave load velocity ratio

F1 . . . F4=Measured heave forces

MMF=Unsprung mass acceleration gain (front)

MMr=Unsprung mass acceleration gain (rear)

DDXu1 . . . DDXu4=Measured unsprung mass accelerations.

The method of isolating a modal force varies depending on whether themeasured corner forces combine positively or negatively to form eachmodal force respectively.

In the case of the heave modal force, as illustrated in FIG. 1, all thesteady state forces and dynamic forces combine positively, so additionterms are used throughout expression (1).

The road input contribution to the heave force is obtained firstly, byadding the corresponding pairs of steady state forces (F₁ +F₂) and(F_(3+F) ₄) measured at the respective load cell of each point ofsupport of the vehicle body. These sums are then each scaled by a commonfactor, 65536.

The dynamic forces are similarly calculated in pairs (DDxu1+DDxu2) and(DDxu3+DDxu4) representing the front and rear wheel/hub accelerations.The front sum is then multiplied by a front mass gain term (MMF) toobtain a force value, and the rear sum by a rear gain term (MMr) toobtain a rear force value.

The front forces are then summed and multiplied by a proportioningfactor (IVrfH) corresponding to the proportion of the heave force whichis attributable to the front of the vehicle. The rear forces aresimilarly summed and multiplied by a proportioning factor (IVrrH)corresponding to the proportion of the heave force attributable to therear of the vehicle.

Clearly, since the isolation of the heave force is effected by themicroprossor 60, and the factors MMF, MMr, IVrFH and IVrrH are notmeasured values, their values may be changed using a suitable inputmeans (not shown) to the microprocessor 60. Changing the four parameterslisted hereinabove therefore alters the calculated value of the heaveforce, and hence may be used to alter the control system's response to aparticular loading condition.

The remaining three modal forces, of pitch, roll and warp, are isolatedby the microprocessor 60 using the following expressions:

For pitch: ##EQU2##

For Roll: ##EQU3##

For Warp ##EQU4## In which: Pf=Generalised pitch force

IVrFP=Inverted pitch load velocity ratio (front)

IVrrP=Inverted pitch load velocity ratio (rear)

Rf=Generalised roll force

IVrFR=Inverted roll load velocity ratio (front)

IVrRR=Inverted roll load velocity ratio (rear)

Wf=Generalised warp force

IVrFW=Inverted warp load velocity ratio (front)

IVrrW=Inverted warp load velocity ratio (rear)

The modal forces are therefore scaled combinations of the measuredvalues, the combinations including either positive or negative values ofthe measured inputs depending on whether, according to the signconvention adopted in FIGS. 1 to 4, the forces combine positively ornegatively as the respective modal forces. Clearly, the control systemof the invention is highly versatile since adjustment of the variousscaling factors introduced in equations 1 to 4, by, for example, key padinput to the microprocessor 60, causes the vehicle to respond to eachmodal force in a predetermined manner. In this way the suspension may bemade, for example, stiff in roll yet soft in heave.

The generalised modal force values given by equations 1 to 4 do not takeaccount of unmeasured loads instigated by the vehicle, such as loadstransmitted to the vehicle body via load paths not including the varioustransducers. Such loads do not give pure modal displacements, andtherefore in taking account of them their effects in cross couplingbetween the modal displacements must also be allowed for.

The type of linkages used commonly to connect the wheel and hubassemblies to the body of a vehicle are not perfect in construction andtherefore part of the load reacted by the unsprung masses is carried bylinkages directly to the vehicle body rather than through the springsand actuators. It is necessary to compensate for these unmeasured loadsin the operation of the active suspension system.

Further, there is envisaged an active suspension system in which loadcells measure only loads on the actuators, without measuring the loadsapplied to the vehicle by the springs. In such a situation, obviously,the forces exerted by the springs are unmeasured loads and hence must beallowed for.

The microprocessor 60 is thus able to produce modified values of themodal forces, as follows: ##EQU5## In which: Hf'=modified generalisedheave force

KHHs=change in heave spring load per unit displacement in the heavedirection

KHPs=change in heave spring load per unit displacement in the pitchdirection

Hx=heave displacement value

Px=pitch displacement value

Pf'=modified generalised pitch force

KPPs=change in load per unit displacement in the pitch direction

KPHs=change in pitch load per unit displacement in the heave direction

Rf'=modified generalised roll force

KRRs=change in roll load per unit displacement in the roll direction

KRWs=change in roll load per unit displacement in the warp direction

Rx=roll displacement value

Wx=warp displacement value

Wf'=modified generalised warp force

WCnx(+/-)=warp acceleration compensation gain

Mxn=scaled longitudinal acceleration

WCDr-yaw acceleration compensation gain in warp

Dr=estimated yaw acceleration

KWWs=change in warp load per unit displacement in the warp direction

KWRs=change in warp load per unit displacement in the roll direction

The modal displacement values (Hx, Px, Rx and Wx) are calculated by themicroprocessor 60 from the measured actuator displacement values (X1 . .. X4). The yaw acceleration estimated value (Dr) is calculated from theyaw rate measurement referred to hereinabove.

The following equations are used by the microprocessor 60 to isolate themodal displacement values: ##EQU6## In which: Vrf=front geometricvelocity ratio multiplied by the gain of the actuator LVIT

Vrr=rear geometric velocity ratio multiplied by the gain of the LVIT

X1 . . . X4=Measured actuator displacements.

The terms KHHs, KPPs, KRRs and KWWs in equations (5) to (8) representthe pure modal forces resulting from the unmeasured vehicle loads andthey are input to the microprocessor as parameters. The values of theparameters may be obtained by testing the vehicle with the controlsystem of the invention switched off.

The values of parameters will depend on the geometry of the suspensionlinkages used. The parameters may be evaluated by executing a series ofmanoeuvres on a smooth surface with a vehicle employing the activesuspension system. Alternatively, an algorithm can be devised toidentify parameters or parameters can be calculated from the precisegeometric arrangement of the suspension linkages.

When the loads applied to the body by the springs are not measureddirectly they become unmeasured loads and must be estimated. The valuesof the co-efficients employed in the estimation of the forcestransmitted by the springs can be evaluated by demanding a smoothlychanging vehicle height relative to the ground for the entire stroke ofeach actuator, the co-efficients being chosen so that the correctedmodal force vector shows no change with actuator position. It isimportant that heave displacement is solely changed during thisprocedure and that the changes take place smoothly and slowly so that novertical inertia forces are generated.

The terms KHPs, KPHs, KRWs and KWRs represent the cross-coupling effectsbetween the modal forces caused by the unmeasured vehicle loads. Fromthese terms, which are also parameters which are input to themicroprocessor 60, it is clear, for example, that the modified rollforce Rf' includes a pure roll mode force (2*KRRs*Rx) caused by theunmeasured vehicle spring loads and force (2*KRWs*Wx) which is crosscoupled from the warp mode, ie it manifests itself in the roll mode as aresult of warp mode displacement due to the unmeasured vehicle springloads. The cross-coupling effect of one particular modal displacement onthe froce associated with another modal direction have been determinedby experiment. Thus, for example, experience has shown that adisplacement of the vehicle body in the roll direction not only changesthe value of the roll load, but also, due to the cross-coupling effectreferred to herein, changes the value of the warp load. Similarconsiderations apply to the other displacement modes, the result beingequations (5) to (8) above.

The terms WXCnx and WCDr indicate that the control system of theinvention takes account of warp modal forces occasioned by thelongitudinal acceleration value (Snx) measured by accelerometer 65 andby the yaw acceleration calculated from the yaw rate measured bygyroscope 66.

The modified modal forces Hf', Pf', Fr' and Wf' may be further processedin microprocessor 60 to produce a demanded output of the actuator 31,which output is implemented by the hydraulic circuit 50, in order tomaintain a constant steady state load on the vehicle sprung mass 20 andtransmit significantly reduced dynamic loads thereto. It istheoretically possible using the control system of the invention totransmit no dynamic load to the vehicle body, but this has the effect ofintroducing zero damping into the suspension system. This is clearlyundesirable from the point of view of vehicle ride, and the controlsystem is therefore adjusted to confer an adequate damping ratio on thesuspension system by transmitting limited dynamic loads to the vehiclebody.

The signal from hub accelerometer 61 may additionally be used to alterthe apparent mass of the wheel/hub assembly. For example, while in theapparatus of FIG. 5 it is always advantageous to determine thecontribution to the load measured by load cell 32 due to theaccelerations of the wheel/hub assembly, it is also possible to processthe signal from accelerometer 61 in an additional feed forward loop theoutput of which is added to the demanded output of the actuator 31. Theoutput of the feed forward loop may be modified to include anyproportion of the accelerometer signal, thereby altering the mass of thewheel/hub assembly apparent at the point where the output of the feedforward loop is added to the demanded output of the actuator 31.

The advantage of processing the signal from accelerometer 61 in thisadditional manner is that the combined signal output to actuator 31 maybe arranged to include not only a proportion of the force measured as aresult of the actual acceleration of the wheel/hub assembly, but also anadditional signal representative of some factor of the mass thereof.This latter signal may, by appropriate programming of the microprocessor60, be adjusted independently of the former signal, so that, forexample, forces due to one apparent wheel/hub assembly mass and anatural frequency of another wheel/hub mass be calculated withinmicroprocessor 60 from a single accelerometer signal measuring thevertical acceleration of a wheel/hub assembly of constant mass.

Such an arrangement is particularly advantageous since it is generallynecessary to take account of the full mass of the wheel/hub assemblyfrom the point of view of loads transmitted to the vehicle body, whileit may be necessary to change the apparent material frequency of thewheel/hub assembly to avoid resonance in the response of system undercertain conditions. If the natural frequency of the wheel/hub assemblycan, as far as an observer of the system from the actuator end isconcerned, be adjusted to be outside the range of output signalfrequencies, the wheel/hub assembly cannot be induced to resonate by anoutput signal of the system.

When a motor vehicle is in motion, clearly both the steady state anddynamic load values vary rapidly, so the microprocessor 60 is capable ofsampling the signals from the various transducers sufficiently rapidlyin sequence to enable the actuator adjustment to be effective incontrolling or correcting the attitude of the vehicle.

The control system of the invention is highly versatile, since any ofthe parameters described herein as inputs to the microprocessor 60 maybe altered. Thus the entire road behaviour of the vehicle may betailored to suit specific requirements and may be adjusted to createdesirable ride and handling conditions of the vehicle at all times ofoperation of the vehicle.

We claim:
 1. A land vehicle suspension system of the kind comprising asuspension linkage connected between a sprung mass of the vehicle and atleast one unsprung mass of the vehicle; actuator means for controllingthe position of the suspension linkage; means for measuring the loadstransmitted to the sprung mass by the actuator means; means forproducing signals proportional to the values of said measured loads;means for modifying said signals to compensate for unmeasured loadsbetween the sprung mass and the at least one unsprung mass acting viaload paths in the suspension linkage which do not include the actuatormeans; and means for controlling the actuator means to apply forces tocontrol the attitude of the vehicle in dependence on said modifiedsignals.
 2. A control system according to claim 1, wherein each unsprungmass of the vehicle is a wheel, hub and tyre assembly, and the sprungmass of the vehicle is substantially the remainder of the mass of thevehicle.
 3. A control system according to claim 1, wherein said meansfor modifying said signals modifies said signals to compensate forportions of said measured loads which do not result from movements ofsaid unsprung mass.
 4. A control system according to claim 1, whereinthe means for modifying said signals includes means for resolving saidsignals as representative of a plurality of modal forces acting on thesprung mass of the vehicle.
 5. A control system according to claim 4,wherein the modal forces are heave, pitch, roll and warp modal forces.6. A control system according to claim 4, wherein the modal values ofthe measured forces are modified by multiplying them by factorsrepresenting the effects of unmeasured loads of the same modesthereupon.
 7. A control system according to claim 4, wherein the modalvalues of the measured forces are modified by multiplying them byfactors representing the effects of unmeasured loads of different modesthereupon.
 8. A control system according to claim 1, including means formeasuring the vertical acceleration of each unsprung mass of thevehicle; means for producing signals proportional to said measuredaccelerations; means for modifying said acceleration proportionalsignals to increase or decrease the extent to which each such signal isdependent on the magnitude of its corresponding unsprung mass; and meansto apply force to control the attitude of the vehicle in dependence onsaid modified acceleration proportional signals.
 9. A control systemaccording to claim 8, wherein said means for modifying said accelerationproportion signals is capable of modifying the signals independently ofother signals in the control system.